Vibration isolation

ABSTRACT

A support for machinery, and for isolating vibration from the machinery, comprises a plurality of mounts, each mount comprising an elastomeric block for completely supporting the static load of the machinery, and active isolation means comprising inertial shakers arranged to maintain essentially a zero stiffness of the mount to excited structural resonances over a desired frequency band above said mount resonant frequency, and to modify the transmission of out of balance forces to the hull. A control system coupled to the inertial shakers includes a means for applying damping force signals, such as to dampen structural resonances, to inhibit the onset of resonant vibration.

CROSS REFERENCE TO RELATED APPLICATIONS

This is a continuation of U.S. patent application Ser. No. 13/292,823,filed Nov. 9, 2011, which is a continuation of U.S. patent applicationSer. No. 11/910,769, filed Oct. 5, 2007, which is a U.S. National Phaseof PCT/EP2006/061407, filed Apr. 6, 2006, which in turn claims priorityto British application no. 0506990.1, filed Apr. 6, 2005, each of whichare incorporated herein in their entirety by reference.

TECHNICAL FIELD

The present invention relates to method and apparatus for isolatingvibrations of machinery from its surroundings, particularly though notexclusively rotating machinery for marine vessels.

BACKGROUND ART

Working machinery is a major source of vibration in marine vessels andconsiderable effort is devoted to developing isolation systems thatreduce transmission to the hull. A particular problem associated withmachinery isolation in marine environments is structural resonance. Thisoccurs principally in the machinery support structure. Such resonanceleads to very high forces transmitted across machinery mounts, and thisposes a very significant vibration problem. Moving machinery generates acomplex spectrum of out-of-balance forces and in marine vesselsconsiderable effort is devoted to developing resilient mounting systemsthat reduce the transmission of these forces across the machinery mountsto the hull.

A common approach to vibration isolation is to mount marine machineryitems on a framework or raft and to support this raft from the hull on aset of rubber mounts. If the supported structures behaved as an idealrigid body, the force transmissibility curve (transfer function fromvibration force to transmitted force) would be as the monotonicallydescending line of FIG. 1. In practice however the supported machineryand its raft will always be flexible to some degree. As a result,structural resonances are excited, and a typical force transmissibilitycurve for resilient mounts is shown by the peaked curve in FIG. 1. Thisillustrates three distinct frequency regimes: the first below the 5 Hzresonance, where the entire force generated by the machinery, primarilythe gravitational force, is transmitted through the resilient mounts;the second, the 5 Hz resonance itself where the machinery, acting as arigid body, is “bouncing” on the resilient mounts, and the third, abovethe 5 Hz resonance, where the machinery is becoming flexible andindividual structural resonant modes are excited. The 5 Hz resonance iscalled the mount resonant frequency. Above this frequency the forcetransmissibility is generally decreasing with increasing frequency andthis results in forces generated by machinery vibrations beingattenuated before arriving at the hull. However, the force attenuationin this regime may be dominated by structural resonances. Structuralresonances act as mechanical amplifiers and hence generate large peaksin the force transmissibility curve as shown.

Because of the compromises that have to be made in designing passiveisolation systems, active and semi-active systems have been proposed. InPatent Application WO 01/18416, and Daley, S., et al, Active vibrationcontrol for marine applications, IFAC Journal Control EngineeringPractice, Volume 12, Number 4, pp 465-474, published 25 Jul. 2003, andin Johnson, A. and Daley, S., A Smart Spring Mounting System for MarineApplications, ISCV11 Conference on Sound and vibration, St Petersburg,July, 2004, an active mount system is proposed comprising an array of alarge number of mounts, each mount comprising an electromagneticactuator in parallel with passive elements to form a composite mount asshown schematically in FIG. 2. In order to avoid transmission of largeforces at frequencies corresponding to supported structure resonances,the mounting system fulfils a number of key requirements. The firstrequirement is for the composite mount not to transmit any additionalforce to the hull as a result of any local displacement of the supportedstructure at its attachment point. As a result no additional force isgenerated on the hull from excited resonances. Thus the composite mountmust have effectively zero stiffness. A second requirement is that tosupport the structure each composite mount must also be able to generatean external demand force for compensating for out of balance forces.Out-of-balance forces, generated by the moving machinery, result in bothlinear and angular displacements of the supported structure. Theexternal demand forces generated by each composite mount are the meanswhereby these linear and angular displacements can be continuouslyopposed to return them towards their equilibrium positions in acontrolled way. As shown in FIG. 2, the actual force on a hull generatedby the composite mount is measured by a load cell (or strain gauge) andcompared with a global demand value, in order continuously to correctthe current of the electromagnet.

In use, each electromagnet is first used to pre-stress the passive mountelements by a static force F so that the change in the force on themachinery may be ±F by increasing or decreasing the current through theelectromagnets; thus a maximum control force of at least 2F is requiredto be generated. When the power is switched off this pre-stress isrelieved. A difficulty with this simple approach is that the largenon-linearity of the electromagnet makes a simple feedback controlunsatisfactory. To overcome this, a more complex local control is neededinvolving both feed-forward of the relative mount displacement andfeedback of the transmitted force

Further improvements in mounting systems are desirable, in particularfor reducing complexity and size of the mounting system.

SUMMARY OF THE INVENTION

The present invention has as one object to provide an improved mount formachinery for isolating machinery vibration.

In a first aspect, the invention provides a method for mountingmachinery, and for isolating vibration therefrom, comprising:

supporting said machinery by means of a support that includes aplurality of resilient mounts, each mount comprising a passive resilientmeans for fully supporting the static load of the machinery, and activeisolation means,

and controlling said active isolation means of each said mount such thatbelow said resonant frequency a performance parameter, primarilystiffness, of each mount is essentially that of the passive resilientmeans, and controlling said active isolation means of each said mountsuch that, above said mount resonant frequency at least over a desiredfrequency band, essentially a zero stiffness of the mount is maintainedto excited structural resonances, and such that rigid body modes ofmovement of said support are compensated, preferably minimised

In a further aspect, the invention comprises a mounting system formounting machinery, and for isolating vibration therefrom, the systemcomprising:

a support for said machinery, including a plurality of resilient mounts,and a control means for controlling the stiffness of the mounts as afunction of frequency,

and each mount comprising a passive resilient means for fully supportingthe static load of the machinery, and active isolation means coupled tosaid control means and wherein the control means is arranged, below theresonant frequency so that a performance parameter, primarily stiffness,of each mount is essentially that of the passive resilient means, andthat, above said mount resonant frequency and over at least desiredfrequency band, the control means is arranged to maintain essentially azero stiffness of the mount to excited structural resonances, and tocompensate, and preferably minimise for rigid body modes of movement ofthe support.

For the purposes of the specification, the “performance” of a mount isdetermined by parameters, primarily stiffness, and to a lesser extent,by its internal damping

The invention realises that there are three main issues for influencingthe maximum control force exerted in a mount system namely:

1) to modify the mount resonance itself the control forces need to becomparable with the gravitational force on the entire machinery raft;2) to maintain “zero stiffness” to local vibrations, particularly thosegenerated by excited resonances;3) to modify and minimise the transmission of out-of-balance forces tothe hull.

As regards 1), it is possible, in accordance with the invention, todesign global control algorithms so that they only modify the mountperformance for frequencies greater than the mount resonant frequency,and that the stiffness, at and below the mount resonant frequencies, canbe made to accurately match that of the passive elements. As aconsequence the static loads remain completely supported by the passiveresilient means alone. This reduces substantially the forces requiredfrom the active isolation means since it now has only to produce forcescomparable with the out-of-balance forces generated by the movingmachinery. These forces are very much less than the static loadsrequired to be carried by the passive elements.

As regards requirements 2) and 3), it has been realised that inaccordance with the invention, only alternating control forces areneeded so that the active actuator elements may be reduced in capacity.This implies that smaller size electromagnetic actuators may beemployed. Alternatively and in accordance with the invention, at leasttwo inertial shakers may be employed, working in opposition to generatezero stiffness. This immediately eliminates the need to pre-stress thepassive element, as is required with systems incorporatingelectromagnets as referred to above, since inertial shakers onlygenerate alternating forces.

The present invention provides in a further aspect a mount for mountingmachinery, and for isolating vibration therefrom, the mount comprising:

a passive resilient means for supporting the static load of themachinery below a mount resonant frequency, and active isolation meanscomprising a plurality of inertial shakers arranged to maintainessentially a zero stiffness of the mount to excited structuralresonances over a desired frequency band above said mount resonantfrequency, and to reduce the transmission of out of balance forces tothe hull.

Inertial shakers have the advantages of being commercially available,significantly smaller, use less power and are inherently much morelinear in their operation than electromagnets. Electrodynamic inertialshakers are well known in the art. Other forms of inertial shakersmaking use of hydraulic, pneumatic, piezoelectric,electro/magneteostrictive drives would also be suitable for thisinvention.

In a typical full-scale marine installation, the forces required fromthese inertial shakers, to offset the out of balance forces, may be arelatively small amount. As regards the requirement to maintain zerostiffness to excited resonances which, because of the large amplitudesthey can generate, could pose greater force demands from the inertialshakers, more powerful hydraulic shakers may be employed as opposed toelectrodynamic shakers.

The inertial shakers in accordance with the invention are coupled tosaid control means for applying appropriate control signals so that theinertial shakers produce required forces to perform their intendedfunction.

The largest force demands on the electromagnets, or inertial shakers,may result from the need to maintain zero stiffness to large amplitudeexcited resonances. However, inertial shakers may be used in a veryforce efficient manner. This can be accomplished by employing selectedinertial shakers to selectively damp any problem resonances—one wouldonly need to extract the energy coupled into the problem resonance fromthe out-of-balance machinery. These damping forces would be no greaterthan the out of balance forces themselves, and generally smaller, sothat only a small increase in the force capability would be requiredfrom the selected inertial The concept of selective damping is disclosedin copending Application WO 01/84012, and British Patent no. 2361757,where damping forces are injected, 90° out of phase with the modaldisplacement, to damp a resonance. The resonance becomes criticallydamped when the energy extracted from it, in this way, is just equal tothe energy coupled into it from the out-of-balance forces so relativelysmall damping forces should suffice.

Thus the present invention provides a much simpler approach than theprior art to the problem of vibration isolation that may give an evenbetter performance with greatly reduced complexity and hence cost. Itmakes use of a relatively small number of controlled mounts that ignorelocal displacements while controlling the response of the machinery'srigid body modes only. This ensures that excited resonances in themounted structure and the machinery do not generate forces on the hull.

In a further aspect, the invention provides a mount for mountingmachinery, and for isolating vibration therefrom, the mount comprising:

a passive resilient means for supporting the static load of themachinery, and a plurality of active isolation elements being disposedaround said passive resilient means, wherein the active isolationelements are inclined to a central axis of the mount so as to compensatefor all rigid body modes of the machinery.

Key features of at least a preferred embodiment of the invention are asfollows.

1) The system is a well-designed passive mounting system in which thepassive mounts have their performance enhanced by the addition of activeelements that can generate an external demand force while maintainingzero stiffness to local displacements. When the active system isswitched off, or in the event of a power failure, the performance fallsto that of the basic passive mounting system—an important “fail-safe”feature.2) The global control algorithms filter out the six rigid body motionsof the supported machinery from the acceleration inputs from an array ofaccelerometers and proximiters. They then generate a set of externaldemand forces to restore these rigid body modes to their equilibriumpositions in a well-controlled way.3) The global control algorithms are designed so that they only modifythe mount performance for frequencies greater than the mount resonantfrequency and the stiffness, at and below the mount resonantfrequencies, accurately matches that of the passive elements. As aconsequence the static loads become supported by the passive elementsalone even when the active control is switched on. Thus the demandforces no longer require a static component and a plurality of inertialshakers can be used in place of electromagnets.4) The global control algorithms can be augmented to detect the onset ofspecific resonances in the supported machinery and to generate selectivedamping forces, 90° out of phase with the modal acceleration, on thisexcited resonance to limit its amplitude. These damping force demandsare fed to selected inertial shakers so that no additional force isgenerated on the hull.5) When the system of the invention is activated all excited structuralresonances are ignored and the force transmissibility, above the mountresonant frequency, can be tailored to fall at a much faster rate andcan include notch filters if required for specific “problem” out ofbalance forces. Again the only forces on the hull are the externaldemand forces to return the rigid body modes to their equilibriumpositions in a well-controlled way.6) By using the active elements to inject disturbance forces, while themachinery is supported by the passive elements alone, one can determinein-situ all the parameters needed by the entire electronic controlsystem, including the detection of specific resonances.

BRIEF DESCRIPTION OF THE DRAWINGS

Preferred embodiments of the invention will now be described withreference to the accompanying drawings, wherein:

FIG. 1 is a graph showing typical force transmissibility for a realstructure (peaked) and an ideal rigid structure (monotonic).

FIG. 2 is a schematic diagram of a prior proposal for an active/passivemounting system for rotating machinery;

FIG. 3 is a block diagram of a control system of the present invention;

FIG. 4 is a schematic internal view of an electrodynamic inertial shakerfor use in the present invention;

FIG. 5 is a perspective view of a first preferred embodiment of a mountaccording to the invention;

FIGS. 6 and 7 are block diagrams of a control system for the activeelements of FIG. 5

FIG. 8 is a perspective view of a second preferred embodiment of a mountaccording to the invention;

FIG. 9 is a schematic perspective view of a mounting system according tothe invention;

FIG. 10 is a graph showing typical force transmissibility, similar toFIG. 1, but in addition showing a curve that may be realised with theinvention;

FIG. 11 is a graph indicating the use of a notch filter to modify thetransmissibility of FIG. 10;

FIG. 12 is a block diagram of the control system of the presentinvention, that augments the system of FIG. 3 for generating dampingforces for damping structural resonances;

FIG. 13 is a block diagram of a control system for an active element ofFIG. 8 to implement the system of FIG. 12;

FIGS. 14 and 15 are graphs of Force Transmissibility for Steel, Rubber &Neoprene; and

FIG. 16 is a graph showing Measured and Predicted Force Transmissibilityfor Rubber.

DESCRIPTION OF THE PREFERRED EMBODIMENT

The motions of a flexible structure can be described as a superpositionof normal modes. These consist of the six zero frequency rigid bodymodes: three translational modes, surge, sway and heave; threerotational modes, roll, pitch and yaw, and the finite frequencystructural resonant modes. If the composite mounts of a mounting systemfor rotating machinery are made to act on the six rigid body modes only,while simultaneously ignoring displacements due to excited resonances,the force transmissibility, for the heave mode, would be as shown by themonotonic descending line curve in FIG. 1. It will be seen that near andbelow the 5 Hz mount resonant frequency the force transmissibility isunchanged but above this frequency there is a major improvement invibration isolation. All the structural resonant peaks, and hence theassociated acoustic signature, have disappeared and the forcetransmissibility is systematically falling at the rate of dB/decade.

In order to achieve this the mounting system according to the inventionmust apply a force that is equal to an external global demand and isindependent of any local displacement (i.e. effectively giving the mount“zero stiffness” to local displacements). This can be achieved by usinglocal controllers, one for each mount. These can ensure a “zerostiffness” to any local vibration, including excited structuralresonances, so that they cannot transmit a force directly through anactuator to the hull. Thus the only force transmitted to the hull is theexternal global demand force.

Referring to FIG. 3, the displacements and velocities of the six rigidbody modes are determined as at 30 by instrumenting a support systemwith an array of accelerometers, and proximiters, each of which canmeasure the local displacement, and hence the local velocity, at itsattachment point. This output data can then be processed, by a matrixtransformation, to determine the displacements and velocities of the sixrigid body modes only. The processing exploits the laws of conservationof linear and angular momentum to filter out, as at 32 the contributionsdue to excited resonances. From the remaining displacements of the sixrigid body modes one can then calculate as at 34, using suitablemathematical models, modal restoring forces and torques for each of thesix rigid body modes, to return them to their equilibrium positions in awell-controlled way. If a standard passive stiffness function is usedfor calculating the rigid body modal restoring forces, then the forcetransmissibility is shown by the descending line in FIG. 1.

Finally one calculates a set of “demand” forces as at 36, one for eachcomposite mount, to generate the required modal forces and torques onthe machinery's six rigid body modes. These demand forces are applied toan array 38 of mounts of the invention for applying restoring forces.

This approach of “zero stiffness” actuators, coupled with a modal globalcontrol law, forms the basis of a mounting system of the invention. Itaims to filter out the effects of resonances at the global observationstage and to use the local controls to generate the required forces onthe rigid body modes while preventing excited structural resonances, orlocal vibrations, from generating forces directly on the hull. Theresult is that the only forces generated on the hull are those needed toreturn the machinery's rigid body modes to their equilibrium positionsin a well-controlled way.

The method employed in the present invention to provide “zero stiffness”is to have an active element in parallel with a passive element and bydesigning its local controller to actively cancel the forces that wouldotherwise be generated by the passive element alone in response to localdisplacements. This local controller must also ensure that the force onthe machinery, as measured by a strain gauge or load cell or otherequivalent force-measuring device, is equal to an external “demand”while the actuator maintains its “zero stiffness” to local vibrations.In particular, it must not transmit any forces directly to the hull froman excited structural resonance.

Referring now to FIG. 9, this shows a mounting system, in accordancewith the invention, for moving (rotating, reciprocating etc) machineryconceptually indicated as at 94. The machinery is mounted on a rigidopen framework raft 96, and the raft is disposed on a rectangular arrayof six mounts 98, each as indicated in FIG. 8. Load cells 64 of eachmount are coupled to a hull structure of a marine vessel (not shown).Marine machinery 94 not only includes the main propulsion units but alsoelectrical generators, lubrication pumps, hydraulic systems, compressedair generators etc. These are commonly all mounted hard mounted on thesingle raft 96. All this machinery, along, with its raft, constitute thestructure supported by the mounts. It is resonances within this entiresupported structure, including the main propulsion machinery itself,that gives the vessel its acoustic signature and which is, in accordancewith the invention, isolated from the hull.

Referring to FIG. 5 showing a perspective view of a mount according tothe invention, a circular raft mount plate 50 is provided for couplingthe mount to raft 96 that supports rotating machinery. The plate iscoupled to a triangular actuator mount plate 52, the apices of whichprovide fixing points for three electromagnets 54. Each electromagnetcomprises an upper mounting plate 56, coupled by means of an armature tothe body 58 of the electromagnet. The base of each body 58 is secured toa further triangular actuator mount plate 60. Plates 52, 60 are securedto a central part of the actuator, comprising a passive element 62formed as a block of elastomer, which is mounted between plates 52, 60.Element 62 is designed to support the weight of the machinery load byitself. The passive element 62 is mounted on a three-axis load cell 64to measure the compression and shear forces generated on the machineryand changes in these forces due to local displacements.

The active elements, that is the electromagnets 54, must be able togenerate a force to cancel the forces that would otherwise be generatedby the passive element alone in response to local displacements. Thisrequires a minimum of three elements arranged as shown, angled towardthe central axis 66 of the mount, to define a tetrahedral configurationwith the axes of the armatures 68 intersecting at an imaginary point 69.

The purpose of the tetrahedral configuration is to be able to generate anet force of a given magnitude and a given direction in space. Thisenables vertical and shear components of the force that would otherwisebe produced by the local passive element to be cancelled—zero stiffness.Finally the net translational forces and torques on the rigid body modesare generated as the sum of the force and from the complete array ofmounts.

The mounting system of the invention shown in FIG. 9 requires fewerelectromagnets than the known system referred to above and, since thetotal mass of the machinery is supported on the passive elements, theactive elements are not required to generate very large forces. Further,if one accepts the performance of the passive system alone forfrequencies at and below the mount resonant frequency, it is possible todesign the global control algorithms to only modify the mountperformance for frequencies greater than the mount resonant frequency,and to accurately match that of the passive elements in below thisfrequency band. This will ensure that at and below the mount resonantfrequency, the performance is controlled by the passive elements alone,while above these frequencies the global control algorithms can bedesigned to modify the mount's vibration isolation performance asrequired. Further, the static loads remain completely supported by thepassive elements alone even when the active control is switched on andthis further reduces the forces required from the active elements. Animportant consequence of this is that the forces generated by the activeelements no longer require a steady component. In order to maintain zerostiffness to local vibrations, including excited resonances, onlyalternating forces, of the appropriate frequency, are required.

The force demands for the active elements are set by two considerations.Firstly there is the need to generate the restoring forces on the rigidbody modes and secondly the need to maintain zero stiffness to excitedresonances. The restoring forces on the rigid body modes will becomparable with the out-of-balance forces generated by the movingmachinery at frequencies greater than the mount resonant frequency. Withreasonably well-balanced machinery these forces can be less than 1/500thof the static force of gravity. Large marine machinery is commonlysupported on an appropriate number of rubber mounts each of 20 tonnecapacity. Thus this would require electromagnets capable of generatingtotal forces up to ±400 newtons [±20,000×10/500]. The actuators in FIG.5 are inclined at 30° to the vertical; each actuator would need togenerate a maximum force of 308 newtons [2×400/(3×cos 30)].

There is also a requirement to maintain zero stiffness to excitedresonances. As indicated in FIG. 1, these can generate large forces onthe hull since the associated mechanical amplification can result inlarge vibration amplitudes at the mounts. If excited resonances dogenerate large amplitude displacements at a mount, larger electromagnetswould be required to maintain zero stiffness.

As regards the control system, shown schematically in FIG. 3, for themount of FIG. 5, the large non-linearity of the electromagnets makes asimple feedback control unsatisfactory. To overcome this, a more complexlocal control is needed as indicated in FIG. 6 and FIG. 7. The strategyinvolves both feed-forward of the relative mount displacement andfeedback of the transmitted force. For small displacements, the mainnon-linearity comes from the behaviour of the electromagnet, however,this is static and an accurate model can be derived. As a result thesystem can be linearised using model inversion techniques. Following aninversion of this type, standard linear methods can be applied to theresidual dynamics in order to meet the local controller objectives. Thelocal controller uses the demand force and the relative displacement togenerate, as accurately as possible, the current demand for theelectromagnet, via a digital switching amplifier, to actively cancel theforce that would otherwise be generated by the passive element alone.The details of the local controller are shown in FIG. 7. The localcontroller also measures the difference between the demand force and themeasured force. This is fed back to a local controller, via a feedbackcompensation unit, to minimise any residual errors.

Where, in accordance with the invention, one accepts the performance ofthe passive system alone, for frequencies at and below the mountresonant frequency, the forces generated by the active elements nolonger require a steady component. This opens up an alternative designfor a mount where the electromagnets of FIG. 5 are replaced by pairs ofelectrodynamic or hydraulic inertial shakers, as shown in FIG. 8, togenerate the equal and opposite forces required to compress or extendthe passive element to give the mount a zero stiffness.

Referring now to FIG. 8, this a perspective view of a second preferredembodiment of a mount according to the invention, where similar parts tothose of FIG. 5 are denoted by the same reference numeral. In FIG. 8,electrodynamic inertial shakers 70 are employed as active elements (adetailed view of the internal construction of one form of such a shakeris shown in FIG. 4, and described below). An upper set of three shakers70 a are mounted on upper mounting plate 52, and a lower set of threeshakers 70 b are mounted on a lower triangular mounting plate 72, so asto so as to oppose the movement of the upper set of shakers, whereby theshakers can exert compressive and tensile forces between plates 54, 72.The axes of the inertial shakers intersect a single point on the axis ofthe mount. Each shaker has a mounting plate 56 coupled to an armature 74that slides in a shaker body 76.

There are three advantages of using inertial shakers in this way.Firstly, there is no longer a need to pre-stress the passive element,since inertial shakers can only generate alternating forces. Secondly,inertial shakers are much more linear in their operation thanelectromagnets so the design of the local controllers becomes simplerand their accuracy better. Thirdly they are more easily installed andare readily available as commercial items.

A schematic form of an electromagnetic electrodynamic inertial shaker isshown in FIG. 4. FIG. 4 shows an electromagnetic electrodynamic activeinertial shaker 40, which comprises a mass 41 consisting of acylindrical permanent magnet whose magnetic axis is vertical. This massis secured by a bolt 44 to two diaphragms 42 which are fixed to thehousing 45. Thus the mass can move up and down in the vertical directionwith the diaphragms 42 acting as springs. The permanent magnet 41 issurrounded by an electrical coil 43, whose axis is also vertical andwhich is attached to the housing 45. When an alternating current ispassed through the coil, the permanent magnet will oscillate verticallythus producing an oscillating vertical inertial force on the housing.The alternating current is provided from the overall control system forthe mount, and generate stiffness functions.

The force demands for inertial shakers are set by the need to generatethe restoring forces on the rigid body modes and secondly the need tomaintain zero stiffness to excited resonances. Returning to the case ofreasonably well-balanced machinery, cited above, these forces can beless than 1/500th of the static force of gravity. Thus to convert a 20tonne capacity passive element to a mount the inertial shakers must becapable of generating total forces of 400 newtons. The mounts in FIG. 8are inclined at 30° to the vertical; each mount would need to generate aforce of 77 newtons [400/(6×cos 30)]. This figure is a quarter of themaximum force of 308 newtons for the electromagnets of FIG. 5.

The more demanding requirement may be the need to maintain zerostiffness to excited resonances. If excited resonances do generate verylarge amplitude displacements at a mount, very much larger forces wouldbe required to maintain zero stiffness. One possibility may be to usesmall hydraulic shakers as these can generate forces of 1,000 newtons,nearly 13 times larger than the figure of 77 newtons required fordealing with out-of-balance forces alone.

The stiffness functions for the shakers are electronically generated soone can use any causal and stabilising function. It is thus possible toimprove the isolation further from that shown by the descending curve inFIG. 1. For example the more steeply descending curve in FIG. 10 showsthe response to a function whereby the high frequency roll-off rate isimproved to dB/decade and the mount resonance is slightly damped. Theother curves are taken from FIG. 1 for comparison. In practice, noise inthe sensors, will limit the maximum performance that can be achieved,but one would expect to achieve a substantially better performance thanthat of FIG. 1.

Clearly the mounting system of the invention requires fewerelectromagnets than the known system referred to above and, since thetotal mass of the machinery is supported on the passive elements, theactive elements are not required to generate very large forces. Further,if one accepts the performance of the passive system alone forfrequencies at and below the mount resonant frequency, it is possible todesign the global control algorithms so that they only modify the mountperformance for frequencies greater than the mount frequency, and toaccurately match that of the passive elements in below this frequencyband. This will ensure that the static loads remain completely supportedby the passive elements alone even when the active control is switchedon and this further reduces the forces required from the activeelements. An important consequence of this is that the forces generatedby the active elements no longer require a steady component. In order tomaintain zero stiffness to local vibrations, including excitedresonances, only alternating forces, of the appropriate frequency, arerequired. This can be achieved by using electrodynamic or hydraulicinertial shakers.

There are three advantages of using inertial shakers in this way.Firstly, there is no longer a need to pre-stress the passive element inorder that changes in the force generated can be either positive ornegative depending on increasing or decreasing the current through anelectromagnet. Secondly, inertial shakers are much more linear in theiroperation than electromagnets so the design of the local controllersbecomes simpler and their accuracy better. Thirdly they are readilyavailable as commercial items.

The array of accelerometers mounted on the machinery for detecting rigidbody motions is also used to detect the onset of a problem resonance,i.e. one that generates large amplitude displacements at one or more ofthe mounts. This merely involves an additional matrix multiplicationwith weighting factors chosen to pick out this resonance. Now the upperset of three inertial shakers 70 a shown in FIG. 8 can also be used toinject damping forces, 90° out of phase with the modal displacement, todamp this resonance, in the manner disclosed in copending Application WO01/84012. The resonance becomes critically damped when the energyextracted from it, in this way, is just equal to the energy coupled intoit from the out-of-balance forces so relatively small damping forcesshould suffice. This additional damping can be added simply as a“software patch” after the problem has been discovered—a furtherpotential cost saving.

It will be noted that electromagnets may not be used for selectivedamping, since if electromagnets attempt to selectively damp excitedresonances they will generate equal but opposite forces on the hull andthe machinery and thus forces on the hull at the resonant frequency.

A control system of the preferred embodiment is shown in FIG. 12 that isan “overlay” of the system of FIG. 3, and specifically directed to theissue of generating appropriate damping forces. Similar parts to thoseof FIG. 3 are denoted by the same reference numeral. FIG. 12 illustratesthe basic principal of selective damping, in this case for damping tworesonances. The acceleration data derived at 30 is processed, by amatrix transformation, to determine the modal velocities as at 100 ofthe problem resonances. From these two modal velocities one cancalculate modal damping forces as at 102 from which one finallycalculates a set of local “demand” forces as at 36, one for each mount,to generate the required modal damping forces

The local controllers for the mounts of FIG. 8 are as shown in FIG. 13.In the absence of a “selective damping control demand force”, the uppermount inertial shakers 70 a and lower mount inertial shakers 70 b aredriven to maintain zero stiffness against local vibrations (relativedisplacements) while maintaining the external force demands to controlthe rigid body motions, as determined by “rigid body control demandforce”. However, the “selective damping control demand force” generatesan additional force, via the above mount shakers only, on the mountedmachinery to damp the specific resonances. The applied raft force is theforce generated on the raft while the transmitted force is the forcetransmitted to the hull. These two forces can differ since the upperinertial shaker can generate an additional damping force that is nottransmitted to the hull.

It is also possible to tailor the force transmissibility to furtherattenuate the transmission of a specific “problem” out-of-balance forceby the use of a “notch filter”. An example of this is shown in FIG. 11where the curve shows a sharp notch filter, centred on 30 Hz, and themonotonic curve is taken from FIG. 10 for comparison. This additionalnotch filter can be added simply as a “software patch” to step 100 ofFIG. 10 after a problem has been encountered—a further potential costsaving.

The mounting system of the invention can be calibrated in situ. One canuse the inertial shakers, to inject disturbance forces while themachinery is supported on the passive elements alone. An analysis of theacceleration responses, along with measurements of the force inputs,makes it possible to derive all the parameters needed by the entireelectronic control system including any additional weighting factors foruse in selectively damping problem resonances.

As regards the construction of the passive element of the mount of theinvention (62 in FIG. 5), measurements were made of the passive forcetransmissibility of three candidates for the passive elements, namelysteel coil springs, lightly damped rubber and more heavily dampedneoprene. The results are shown in FIGS. 14 and 15. It will be seen fromFIG. 11 that the steel coil springs have the lowest internal damping andhence the largest mount resonance. In the case of steel springs, FIG. 15not only displays the mount resonance (around 10 Hz) but also higherfrequency spring resonances at about 150, 340 and 360 Hz. The absence ofresonances in rubber and neoprene results in much lower forcetransmissibilities at frequencies above 150 Hz. Below −60 dB the signalsfall beneath the noise floor of the instrumentation. In all cases thereis a resonance at around 250 Hz. This is a resonance in the load cellthat causes the force transmissibility to rise. However, the roll-offrate, above 250 Hz, is significantly improved, in the case of rubber andneoprene, due to the additional high frequency isolation produced bythis resonance—it acts as a double mounting system. In the case of steelsprings this improvement is masked by resonances in this region. It isclear that the use of elastomers results in a better overall performancethan steel coil springs. Natural rubber has a greater roll-off rate inthe region immediately above the mount resonance and is often preferredin marine environments due to its superior tear strength. However, thecritical issue is how well the response of the three candidates can bemodelled so that the actuator can be controlled accurately. A test ofthe modelling accuracy is shown by a comparison of the measured andpredicted force outputs for a random excitation. FIG. 16 shows thiscomparison for the rubber passive elements. The modelling accuracy forsteel springs and neoprene were broadly similar but their detailedprecision was less good, particularly in the case of the steel springswhere the errors were greatest near the spring's resonant frequencies.

1. A mounting system for mounting machinery, and for isolating vibrationtherefrom, the system comprising: a support for said machinery,including a plurality of resilient mounts, and a control means forcontrolling the stiffness of the mounts as a function of frequency, eachmount comprising a passive resilient means for fully supporting thestatic load of the machinery, and active isolation means, wherein saidcontrol means includes a notch filter for attenuating the transmissionof a specific out-of-balance force.
 2. A system according to claim 1,wherein said active isolation means of each mount comprises a pluralityof active isolation components.
 3. A system according to claim 2,wherein said components comprise inertial shakers.
 4. A system accordingto claim 3, wherein each component comprises a plurality of inertialshakers, one shaker being disposed in an opposing direction to anothershaker to form a pair.
 5. A system according to claim 2, wherein eachsaid component comprises an electromagnet.
 6. A system according toclaim 2, wherein said components are inclined at an angle to a centralaxis of their respective resilient mount.
 7. A system according to claim1, wherein the active isolation means comprises a plurality of inertialshakers, disposed relative to one another and arranged to maintainessentially a zero stiffness of the mount to excited structuralresonances over a desired frequency band above said mount resonantfrequency, and to modify the transmission of out of balance forces to ahull.
 8. A system according to claim 1, wherein said support comprise araft, with said plurality of resilient mounts forming an array mountingthe raft to a structure.
 9. A system according to claim 1, wherein thecontrol system includes means for sensing motion of the support in rigidbody modes of motion, means for filtering, from the detected motion,displacement arising from excited resonances, means for calculatingmodal restoring forces, and means for applying local demand forces toeach said resilient mount for compensating for rigid body modes ofmotion.
 10. A system according to claim 1, said control means includingmeans for applying a damping force to said resilient mounts to inhibitdevelopment of predetermined structural resonances.
 11. A systemaccording to claim 1, wherein said notch filter is a notch filter forfiltering predetermined structural resonances.
 12. A system according toclaim 1, each mount including a three-axis force measuring means formeasuring transmitted force.
 13. A mounting system for mountingmachinery, and for isolating vibration therefrom, the system comprising:a support for said machinery, including a plurality of resilient mounts,and a controller for controlling the stiffness of the mounts as afunction of frequency, each mount comprising a passive resilient elementfor fully supporting the static load of the machinery, and an activeisolation element, wherein said controller includes a notch filter forattenuating the transmission of a specific out-of-balance force.
 14. Asystem according to claim 13, wherein said active isolation element ofeach mount comprises a plurality of active isolation components.
 15. Asystem according to claim 14, wherein each said component comprises anelectromagnet.
 16. A system according to claim 14, wherein saidcomponents are inclined at an angle to a central axis of theirrespective resilient mount.
 17. A system according to claim 13, whereinsaid support comprise a raft, with said plurality of resilient mountsforming an array mounting the raft to a structure.
 18. A systemaccording to claim 13, said controller including a force applyingelement for applying a damping force to said resilient mounts to inhibitdevelopment of predetermined structural resonances.
 19. A systemaccording to claim 13, wherein said notch filter is a notch filter forfiltering predetermined structural resonances.
 20. A system according toclaim 13, each mount including a three-axis force measuring means formeasuring transmitted force.